Internal combustion engines have a cylinder block and at least one cylinder head, which are connected to one another to form the at least one cylinder. To accommodate the pistons and the cylinder liners, the cylinder block has cylinder bores. The pistons are guided in an axially movable manner in the cylinder liners and, together with the cylinder liners and the at least one cylinder head, form the combustion chambers of the internal combustion engine.
Modern internal combustion engines are operated almost exclusively by means of a four stroke operating method. The charge cycle includes the expulsion of the combustion gases via the at least one exhaust port of the at least one cylinder in order to discharge the exhaust gases via the exhaust gas discharge system and filling with fresh mixture or charge air via the intake system by way of the at least one intake port. In order to control the charge cycle, an internal combustion engine requires control elements and actuating devices for actuating said control elements. To control the charge cycle, the control elements used in four stroke engines are almost exclusively reciprocating valves, which perform an oscillating stroke motion during the operation of the internal combustion engine and, in this way, open and close the intake and exhaust ports. The valve actuating mechanism required for the movement of the valves, including the valves themselves, is referred to as a valve train.
The at least one cylinder head is generally used to accommodate at least some portion of this valve train, which are covered by means of a cover, with the cover being connected (preferably removably) to the cylinder head and surrounding the parts of the valve train in the manner of a housing.
For actuation of a valve, use is made, on the one hand, of valve spring means in order to preload the valve in the direction of the valve closing position, i.e. to subject it to a preloading force, and, on the other hand, of valve actuating devices in order to open the valve counter to the preloading force of the valve spring means. A valve actuating device comprises a camshaft, on which a multiplicity of cams is arranged and to which rotation about the longitudinal axis thereof is imparted in such a way by the crankshaft—by means of a chain drive, for example—that the camshaft and, together with the latter, the cams revolves or revolve at half the speed of the crankshaft.
A fundamental distinction is drawn between a low-mounted camshaft and an overhead camshaft. In this context, the reference point is the parting plane between the cylinder head and the cylinder block. If the camshaft is above this parting plane, it is an overhead camshaft and otherwise a low-mounted camshaft. Overhead camshafts are generally accommodated by the cylinder head, i.e. on the opposite side of the cylinder head from the parting plane.
It is the object of the valve train to open and close the intake and exhaust ports of the cylinders at the right times, the aim being rapid opening of as large as possible flow cross sections in order to minimize throttling losses in the inflowing and outflowing gas flows and to ensure optimum filling of the combustion chambers with fresh mixture and effective, i.e. complete, removal of the exhaust gases.
The intake lines, which lead to the intake ports, and the exhaust lines, which are connected to the exhaust ports, are at least partially integrated into the cylinder head. The exhaust lines of the cylinders are brought together to form a common overall exhaust line or brought together in groups to form a plurality of overall exhaust lines, thereby forming an exhaust gas discharge system. In general, the intake lines of the intake system are fed, i.e. supplied with fresh mixture or charge air, from a common overall intake line.
Internal combustion engines are being fitted more and more frequently with a pressure charging system, pressure charging being primarily a method for boosting power, in which the charge air required for the combustion process in the engine is compressed, thus enabling a larger mass of charge air to be fed to each cylinder in each working cycle. It is thereby possible to increase the fuel mass and hence the mean pressure.
In general, an exhaust gas turbocharger is used for pressure charging, in which a compressor and a turbine are arranged on the same shaft, the hot exhaust gas flow being fed to the turbine, expanding in said turbine and releasing energy in the process, and thereby imparting rotation to the shaft. The energy released by the exhaust gas flow to the shaft is used to drive the compressor, which is likewise arranged on the shaft. The compressor conveys and compresses the charge air fed to it, and pressure charging of the cylinders is thereby achieved.
As in the case of the internal combustion engine according to the invention, a charge air cooler is preferably provided in the intake system, by means of which cooler the compressed charge air is cooled before entry to the at least one cylinder. The cooler lowers the temperature and thus increases the density of the charge air, and the cooler therefore also contributes to better filling of the at least one cylinder, i.e. to a larger air mass. Cooling results in increased density.
A coolant-fed charge air cooler is preferably used, comprising heat-exchange elements through which coolant flows, between which the charge air flows and is cooled.
In order to comply with future limits for pollutant emissions, and especially to reduce nitrogen oxide emissions, exhaust gas recirculation (EGR) is often employed, in which combustion gases are taken from the exhaust gas discharge system on the exhaust side and fed back into the intake system on the intake side.
With an increasing exhaust gas recirculation rate, nitrogen oxide emissions can be significantly reduced. The exhaust gas recirculation rate xEGR is determined by xEGR=mEGR/(mEGR+mfresh air), where mEGR denotes the mass of exhaust gas fed back and mfresh air denotes the fresh air fed in. In order to achieve a significant reduction in nitrogen oxide emissions, high exhaust gas recirculation rates are required, and these can be of the order of xEGR≈60% to 70%.
During the operation of an internal combustion engine with turbocharging and simultaneous use of exhaust gas recirculation, a conflict can arise if the recirculated exhaust gas is taken from the exhaust gas discharge system upstream of the turbine by means of high-pressure EGR and is no longer available to drive the turbine.
As the exhaust gas recirculation rate is increased, the exhaust gas flow introduced into the turbine simultaneously decreases. The reduced exhaust gas mass flow through the turbine entails a lower turbine pressure ratio and, as a result, the boost pressure ratio likewise decreases, and this is equivalent to a lower compressor mass flow. Apart from the decrease in boost pressure, additional problems can arise in the operation of the compressor as regards the pulsation limit.
For this reason, there is a need for concepts which ensure sufficiently high boost pressures with simultaneously high exhaust gas recirculation rates—especially in the part load range. What is termed “low-pressure EGR” offers one approach to a solution.
In contrast to the high-pressure EGR system already mentioned, which takes exhaust gas from the exhaust gas discharge system upstream of the turbine and introduces it into the intake system downstream of the compressor, the procedure in the case of low-pressure EGR is to feed exhaust gas back to the intake side after it has already flowed through the turbine. For this purpose, the low-pressure EGR system comprises a recirculation line, which branches off from the exhaust gas discharge system downstream of the turbine and opens into the intake system upstream of the compressor.
The exhaust gas fed back to the intake side by means of low-pressure EGR is mixed with fresh air upstream of the compressor. The mixture of fresh air and recirculated exhaust gas produced in this way forms the charge air which is fed to the compressor and compressed, with the compressed charge air being cooled downstream of the compressor in the charge air cooler.
The fact that the exhaust gas is passed through the compressor as part of low-pressure EGR does not have a negative effect since use is generally made of exhaust gas which has been subjected to exhaust gas aftertreatment, more particularly in the particle filter, downstream of the turbine. There is therefore no need to fear deposits in the compressor that change the geometry of the compressor, in particular the flow cross sections, and in this way impair the efficiency of the compressor. For the same reasons, moreover, the untreated exhaust gas recirculated by means of high-pressure EGR is not passed through the charge air cooler since this would lead to contamination and deposits in the cooler.
On the other hand, problems can arise downstream of the compressor owing to the cooling of the compressed charge air. According to the prior art, the charge air cooler is often arranged to the side of and adjacent to the internal combustion engine, e.g. at the level of the crank case, i.e. at the level of the cylinder block or the oil sump. This is also done with a view to dense packaging of the overall drive unit.
During cooling, liquids which were previously still present in gaseous form in the charge air, in particular water, can condense out if the dew point of one component of the gaseous charge air flow is undershot. Owing to the arrangement of the charge air cooler and the fact that the precipitated condensate is not removed continuously from the charge air flow and fed to the cylinders in very small quantities due to the kinetics, condensate may collect in the charge air cooler and may then be introduced into the intake system from the charge air cooler unpredictably all of a sudden in relatively large quantities, e.g. in the case of transverse acceleration due to cornering, a hill or a bump. The latter case is also referred to as water hammer, which may lead not only to severe disruption of the operation of the internal combustion engine but also may damage engine components downstream of the cooler.
The problems described above become more severe as the recirculation rate increases since the proportions of the individual exhaust gas components in the charge air, in particular that of water in the exhaust gas, inevitably increase with the increase in the quantity of exhaust gas recirculated. According to the prior art, the quantity of exhaust gas recirculated by means of low-pressure EGR is therefore limited in order to reduce the quantity of water that condenses out or to prevent water from condensing out. The required limitation of low-pressure EGR, on the one hand, and the high exhaust gas recirculation rates required for a significant reduction in nitrogen oxide emissions, on the other hand, lead to differing aims in the dimensioning of the recirculated quantity of exhaust gas.
The indicated conflict in the case of an internal combustion engine pressure charged by means of exhaust gas turbocharging and fitted with a low-pressure EGR system and a charge air cooling system cannot be resolved by the prior art.
According to the prior art, the required high recirculation rates can only be achieved by means of high-pressure EGR, and the disadvantages associated therewith have to be accepted. Consequently, the advantages of low-pressure EGR can only be exploited to a limited extent.
Arranging the compressor on the exhaust side entails a comparatively long travel distance between the compressor and the intake port on the cylinder. However, this travel distance should be kept as short as possible in order to ensure rapid turbocharger response and to minimize the pressure loss in the charge air flow due, in particular, to deflections. Moreover, a long travel distance in the intake system has disadvantages in terms of the noise behavior of the internal combustion engine and leads to low-frequency noise emissions.